Compressors consist of a series of rotating and stationary blade rows in which the combination of a rotor (circular rotating blade row) and a stator (circular stationary blade row) forms one stage. Inside the rotor, kinetic energy is transferred to the gas flow (usually air) by the individual airfoil blades. In the following stator, this energy is manifested as a pressure rise in the gaseous air as a consequence of deceleration of the gaseous air flow. This deceleration of the gaseous air flow is induced as a result of the design of the stator section. The pressure ratio (exit pressure/inlet pressure) of a single stage is limited because of intrinsic aerodynamic factors, so several stages are connected together in many turbo compressors to achieve higher pressure ratios than could be achieved by a single stage.
The maximum achievable pressure ratio of a turbo compressor is established by the so-called stability limit of the compressor given by the characteristic of the compressor and the gaseous air flowing through the compressor at any time. As the pressure in the compressor increases, the aerodynamic loading on the compressor blades must also increase. At full speed operation of a multi stage compressor, the rear stages carry the majority of the aerodynamic load (and attendant stress), and the stability limit is established by the limits inherent in the design of these stages. When operating at lower speeds, the stability limit of the compressor is established by limitations deriving from characteristics related to the front stages of the compressor.
In the normal stable working range of a compressor stage, axial flow of gaseous air through all of the vane channels between the compressor blades takes place equally and continuously as the air volume is transported through the channels. However, a compressor stage can also operate in a state known as an unstable working range. In this unstable working range, a stall condition can be present in the interaction between the air flow and the airfoil blades which can contribute to substantial variations in the internal pressure profile of the compressor. These pressure variations can, in turn, cause substantial stress to the blades of the compressor. Ultimately, this stress can damage the blades if the compressor continues to operate in the unstable working range for any length of time. Operation in the unstable working range is inefficient at best and potentially destructive; this mode of operation should be avoided as much as possible.
The development of a stall in a stage of the compressor proceeds from the interaction of individual airfoil blades with the gaseous air flowing through the vanes associated with those individual blades. Ideally, the gaseous air fluid flow should be axially continuous through the compressor; however, high blade loads can induce localized disruptions to that continuous flow.
The air fluid flow around each blade has an associated flow boundary layer which covers each blade and coheres to the blade. The flow boundary layer associated with a rotor blade will rotate as an associated entity of the blade as the blade itself rotates. At the downstream edge of each blade, this flow boundary layer melds into an associated flow boundary entity known as, alternatively, the Dellenregion, wake region, or delve region which is characterized by a localized reduction in both pressure and flow velocity. With increasing load, this wake region correspondingly will extend until a critical mass or size is achieved; when the wake region on the downstream edge of the blade achieves this critical size, it fractures or fragments into (1) a (new) smaller wake region which is still coherent with the blade and (2) a "flow boundary layer part" which physically separates from the wake region. Studies have indicated that these "flow boundary layer parts", separated from the rotor blades, move radially outwards from the axis of rotation due to centrifugal forces and collect at the inner circumferential surface of the compressor housing. This collection of separated flow boundary layer parts "swirls" and effectively establishes a turbulent fluid layer (or collection of swirled separated regions) at the inner surface of the housing; this turbulent fluid layer has associated stochastic pressure fluctuations which are useful in the present invention. For the purpose of this disclosure, this initial state associated with an increasing compressor loading will be termed as a "separated flow pre-stall".
With further increasing load, disrupted flow zones downstream of the blades expand in size and/or increase in number. Disruption of the continuous air flow through either groups of non-contiguous single-blade channels or whole sections of contiguous blade channels may occur. This blockage may be characterized as a sort of "bubble-like" entity which, in general, moves circumferentially throughout the stage with a rotational speed up to 0.5 times the rotor frequency. This phenomenon is known as "rotating stall". In stages with large blade heights, only the radially outer part of the blade channels is blocked and this situation is known as a "full span stall". With increasing load, the entire set of blade channels in a stage can be effectively blocked, resulting in an event and condition known as a "full span stall". In case of compressor stages having small overall diameters, "full span stall" can occur directly without transition through "part span stall" status.
Another phenomenon, which may derive from rotating stall or also may occur suddenly with increased blade loading, is the "compressor surge". In this state, the whole circumference of one stage (usually the last one) has stalled (full span stall in the full blading). Then, the compressor cannot work any longer against the back-pressure of this one stage and the flow in the compressor breaks down. The high pressure gas flows back from the outlet to the compressor intake until the pressure at the compressor outlet is reduced enough so that a moderate blade load allows normal working again. When the back pressure is not reduced, this changing operation will be continued. These fluctuations will take place with very low frequencies (typically a few Hertz) and will destroy the compressor within a short time of operation because the rotor is respectively shifted axially fore and aft. Furthermore, the compressor surge will be accompanied by fluctuations in the continuous overall air flow to the firing chamber in case of a gas turbine; these fluctuations can disrupt the environment in the firing chamber of the turbine in such a manner as to extinguish the "flame" in the firing chamber or (in some rare instances) establish the prerequisite environment for a backfire of the turbine through the compressor. A compressor should not be operated under such conditions; at best, operation will be inefficient for those stages wherein stall effects occur.
On the other hand, it is desirable to operate a compressor in an optimally efficient manner (that is as close as possible to the appropriate maximum obtainable mass flow rate given by the overall status of the compressor). Contemporary turbo engines are usually equipped with fuel or energy control systems which measure and output a variety of operating parameters for the overall engine. Included in such control systems are highly accurate pressure sensing devices or systems. For example, a pressure measuring system is described in PCT Publication (with International Publication Number WO 94/03785 filed Jun. 16, 1994 and published Feb. 17, 1994) titled ADAPTOR FOR MOUNTING A PRESSURE SENSOR TO A GAS TURBINE HOUSING. This publication shows a preferred pressure measuring system for use in the invention. Material from this publication is also presented with respect to FIG. 7, FIG. 8, and FIG. 9. Other examples of pressure measuring systems are described in U.S. Pat. No. 4,322,977 entitled "Pressure Measuring System", filed May 27, 1980 in the names of Robert. C. Shell, et al; U.S. Pat. No. 4,434,664 issued Mar. 6, 1984, entitled "Pressure Ratio Measurement System", in the names of Frank J. Antonazzi, et al.; U.S. Pat. No. 4,422,335 issued Dec. 27, 1983, entitled "Pressure Transducer" to Ohnesorge, et al.; U.S. Pat. No. 4,449,409, issued May 22, 1984, entitled "Pressure Measurement System With A Constant Settlement Time", in the name of Frank J. Antonazzi; U.S. Pat. No. 4,457,179, issued Jul. 3, 1984, entitled "Differential Pressure Measuring System", in the names of Frank J. Antonazzi, et al.; and U.S. Pat. No. 4,422,125 issued Dec. 20, 1983, entitled "Pressure Transducer With An Invariable Reference Capacitor", in the names of Frank J. Antonazzi, et al.
U.S. Pat. No. 4,216,672 to Henry et al, discloses an apparatus for detecting and indicating the occurrence of a gas turbine engine stall which operated by sensing sudden changes in a selected engine pressure. A visual indication is also provided.
U.S. Pat. No. 4,055,994 to Roslyng et al discloses a method and a device of detecting the stall condition of an axial flow fan or compressor. The method and device measure the pressure difference between the total air pressure acting in a direction opposite to the direction of the revolution of the fan wheel and a reference pressure corresponding to the static pressure at the wall of a duct in substantially the same radial plane.
U.S. Pat. No. 4,618,856 to Frank J. Antonazzi discloses a detector for measuring pressure and detecting a pressure surge in the compressor of a turbine engine. The detector is incorporated in an analog to a digital pressure measuring system which includes a capacitive sensing capacitor and a substantially invariable reference capacitor.
While a wide variety of pressure measuring devices can be used in conjunction with the present invention, the disclosures of the above-identified patents and the articles mentioned next are hereby expressly incorporated by reference herein for a full and complete understanding of the operation of the invention.
The article "Rotating Waves as a Stall Inception Indication in Axial Compressors" of V. H. Garnier, A. H. Epstein, E. M. Greitzer as presented at the "Gas Turbine and Aeroengine Congress and Exposition" from Jun. 11 to 14, 1990, Brussels, Belgium, ASME Paper No. 90-GT-156, discloses the observation of rotating stall. In case of a low speed compressor, the axial velocity of air flow is measured by several hot wire anemometers distributed around the circumference of the compressor. From the respective sensor signals, complex Fourier coefficients are calculated, which coefficients contain detailed information on the wave position and amplitude as a function of time of a wave traveling along the circumference of the compressor. These traveling waves are to be identified with rotating stall waves. In case of a high speed compressor, several wall mounted, high-response, static pressure transducers are employed, from which sensor signals first and second Fourier coefficients are being derived. However, this direct spectral approach does not directly yield information on compressor stability, since the height of the rotating stall wave peak is a function of both the damping of the system and the amplitude of the excitation. To estimate the wave damping, a damping model is fitted to the data for an early time estimate of the damping factor. By this technique, a rather short warning time may be available (in the region of tens to hundreds of rotor revolutions) to take corrective action (changing the fuel flow, nozzle area, vane settings etc.) to avoid compressor surge.
In the article "Fast Response Wall Pressure Measurement as a Means of Gas Turbine Blade Fault Identification" of K. Nathioudakis, A. Papathanasious, E. Loukis and L. Papailiou, as presented at the "Gas Turbine and Aeroengine Congress and Exposition" at Brussels, Belgium, from Jun. 11-14, 1990, ASME Paper No. 90-GT-341, it is mentioned that rotating stall is accompanied by the appearance of distinct waveforms in the measured pressure, corresponding to a rotational speed which is a fraction of the shaft rotational speed.
The systems known in the art cannot detect an unstable operating condition based on the preliminary indications of instability. They can only detect well established unstable conditions in an advanced state and, therefore, must avoid operation in the region where damage could result to the compressor from the more subtle kinds of instability. In order to avoid operation in the region where damage could result to the compressor, prior art compressor control systems operate with a high safety margin; this margin is well below the maximum possible mass flow rate of the compressor. In effect, the prior art compressor must therefore operate in a less efficient and a less economical mode than be realized with the subject of this invention.
Furthermore, the prior art control systems can detect an existing tendency of the compressor towards a stall condition or a surge condition only at a very short time before the actual occurrence of stall or surge. In many cases there is not enough time left after the above detection to take corrective actions for avoiding stall or surge.
The reduction of the risk of compressor stall and compressor surge is a further reason for the prior art compressor control systems to operate with the high safety margin.